US10508709B2 - Vibration damping device and method for designing the same - Google Patents

Vibration damping device and method for designing the same Download PDF

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Publication number
US10508709B2
US10508709B2 US16/321,554 US201716321554A US10508709B2 US 10508709 B2 US10508709 B2 US 10508709B2 US 201716321554 A US201716321554 A US 201716321554A US 10508709 B2 US10508709 B2 US 10508709B2
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Prior art keywords
vibration damping
damping device
restoring force
force generating
mass body
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US20190226552A1 (en
Inventor
Yoichi OI
Masaki Wajima
Hiroki Nagai
Takao Sakamoto
Takahiro Ryu
Takashi Nakae
Kenichiro Matsuzaki
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NATIONAL UNIVERSITY Corp OITA UNIVERSITY
Aisin AW Co Ltd
Kagoshima University NUC
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NATIONAL UNIVERSITY Corp OITA UNIVERSITY
Aisin AW Co Ltd
Kagoshima University NUC
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Assigned to KAGOSHIMA UNIVERSITY, NATIONAL UNIVERSITY CORPORATION OITA UNIVERSITY, AISIN AW CO., LTD. reassignment KAGOSHIMA UNIVERSITY ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: NAKAE, TAKASHI, RYU, TAKAHIRO, MATSUZAKI, KENICHIRO, NAGAI, HIROKI, OI, Yoichi, SAKAMOTO, TAKAO, WAJIMA, Masaki
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/14—Suppression of vibrations in rotating systems by making use of members moving with the system using masses freely rotating with the system, i.e. uninvolved in transmitting driveline torque, e.g. rotative dynamic dampers
    • F16F15/1407—Suppression of vibrations in rotating systems by making use of members moving with the system using masses freely rotating with the system, i.e. uninvolved in transmitting driveline torque, e.g. rotative dynamic dampers the rotation being limited with respect to the driving means
    • F16F15/145—Masses mounted with play with respect to driving means thus enabling free movement over a limited range
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/12—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon
    • F16F15/1202—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the damping action being at least partially controlled by centrifugal masses
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/12—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon
    • F16F15/131—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses
    • F16F15/13128—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses the damping action being at least partially controlled by centrifugal masses
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/12—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon
    • F16F15/131—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses
    • F16F15/133—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses using springs as elastic members, e.g. metallic springs
    • F16F15/137—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses using springs as elastic members, e.g. metallic springs the elastic members consisting of two or more springs of different kinds, e.g. elastomeric members and wound springs
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/14—Suppression of vibrations in rotating systems by making use of members moving with the system using masses freely rotating with the system, i.e. uninvolved in transmitting driveline torque, e.g. rotative dynamic dampers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H45/00—Combinations of fluid gearings for conveying rotary motion with couplings or clutches
    • F16H45/02—Combinations of fluid gearings for conveying rotary motion with couplings or clutches with mechanical clutches for bridging a fluid gearing of the hydrokinetic type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F2228/00—Functional characteristics, e.g. variability, frequency-dependence
    • F16F2228/001—Specific functional characteristics in numerical form or in the form of equations

Definitions

  • vibration damping devices including a restoring force generating member that can swing as a support member rotates and an inertial mass body that is coupled to the support member via the restoring force generating member and swings with motion of the restoring force generating member as the support member rotates, and methods for designing the same.
  • a vibration damping device including a flywheel mass body that receives a centrifugal force and functions as a restoring force generating member and an annular inertial mass body coupled to the flywheel mass body via a connecting rod is known as this type of vibration damping device (see, e.g., Patent Document 1).
  • the inertial mass body swings with the swinging motion of the flywheel mass body, and vibration of the support member can be dampened by vibration that is transmitted from the inertial mass body to the support member.
  • a centrifugal pendulum rotational speed-adaptive dynamic vibration absorber including a support member coupled to a rotary element that is rotated by power from a drive device and mass bodies that are coupled to the support member and swing about a pendulum fulcrum is also known as a vibration damping device (see, e.g., Patent Document 2).
  • This rotational speed-adaptive dynamic vibration absorber is designed to have an effective order qeff that is larger than the order q of excitation of a drive device by a predetermined order offset value qF in view of the influence of oil.
  • Patent Document 1 German Patent Application Publication No. 102012212854
  • Patent Document 2 Published Japanese Translation of PCT Application No. 2011-504987 (JP 2011-504987 A)
  • the effective order qeff is made larger than the excitation order q of the drive device by the predetermined order offset value qF in view of resistance (viscous resistance) to relative motion of the mass bodies with respect to rotating oil in order to make the order of vibration that can be most satisfactorily damped by the dynamic vibration absorber equal to the excitation order q.
  • centrifugal pendulum vibration damping devices can basically have satisfactorily vibration damping capability.
  • vibration damping devices including such an inertial mass body as described in Patent Document 1 may not have satisfactorily vibration damping capability even if the order of vibration that can be most satisfactorily damped by the vibration damping device is made equal to the excitation order of a drive device. There is still room for improvement in terms of more properly setting the order to improve vibration damping capability.
  • a vibration damping device of the present disclosure is a vibration damping device including a support member that rotates, together with a rotary element to which torque from an engine is transferred, about a center of rotation of the rotary element, a restoring force generating member that is coupled to the support member and that is swingable as the support member rotates, and an inertial mass body that is coupled to the support member via the restoring force generating member and that swings, with motion of the restoring force generating member, about the center of rotation as the support member rotates, wherein moments of inertia J 1 , J 2 , mass m, and distances L 3 and L 4 are determined so that torque fluctuation of an object for which vibration is to be damped, which is derived based on angular displacement ⁇ and angles ⁇ and ⁇ obtained by solving the following expressions (A) and (B), is equal to a target value
  • vibration damping capability can be improved by making a reference order, namely a value to which an order of the vibration damping device converges as amplitude of vibration of input torque decreases, larger than an excitation order of an engine rather than by making the reference order equal to the excitation order.
  • the moments of inertia J 1 , J 2 , the mass m, and the distances L 3 and L 4 are determined so that the torque fluctuation of the object for which vibration is to be damped, which is derived based on the angular displacement ⁇ and the angles ⁇ and ⁇ obtained by solving the expressions (A) and (B), is equal to the target value.
  • the reference order can thus be made to fall within a proper range that is larger than the excitation order of the engine, whereby the vibration damping capability is further improved.
  • FIG. 1 is a schematic configuration diagram of a starting device including a vibration damping device of the present disclosure.
  • FIG. 2 is a sectional view of the starting device shown in FIG. 1 .
  • FIG. 3 is a front view of the vibration damping device of the present disclosure.
  • FIG. 4 is an enlarged sectional view of a main part of the vibration damping device of the present disclosure.
  • FIG. 5A is a schematic view illustrating operation of the vibration damping device of the present disclosure.
  • FIG. 5B is a schematic view illustrating operation of the vibration damping device of the present disclosure.
  • FIG. 5C is a schematic view illustrating operation of the vibration damping device of the present disclosure.
  • FIG. 6 is a graph illustrating the relationship between the engine speed and the torque fluctuation T Fluc in an output element of a damper device of the present disclosure.
  • FIG. 7 is a graph illustrating the relationship between the reference order q ref and the torque fluctuation T Fluc in the output element of the damper device of the present disclosure.
  • FIG. 8 is a flowchart showing an example of a procedure for designing the vibration damping device of the present disclosure.
  • FIG. 9 is a flowchart showing another example of the procedure for designing the vibration damping device of the present disclosure.
  • FIG. 10 is a schematic view illustrating another vibration damping device of the present disclosure.
  • FIG. 11 is a schematic view illustrating a further vibration damping device of the present disclosure.
  • FIG. 12 is a front view of a further vibration damping device of the present disclosure.
  • FIG. 13 is an enlarged view of a further vibration damping device of the present disclosure.
  • FIG. 14 is an enlarged sectional view of a main part of the vibration damping device shown in FIG. 13 .
  • FIG. 15 is an enlarged sectional view of a main part of a modification of the vibration damping device shown in FIG. 13 .
  • FIG. 16 is an enlarged view of a main part of another modification of the vibration damping device shown in FIG. 13 .
  • FIG. 17 is an enlarged view of a main part of a modification of the vibration damping device shown in FIG. 13 .
  • FIG. 18 is a schematic configuration diagram of a modification of the damper device including the vibration damping device of the present disclosure.
  • FIG. 19 is a schematic configuration diagram of another modification of the damper device including the vibration damping device of the present disclosure.
  • FIG. 1 is a schematic configuration diagram of a starting device 1 including a vibration damping device 20 of the present disclosure.
  • the starting device 1 shown in the figure is mounted on, e.g., a vehicle including an engine (internal combustion engine) EG serving as a drive device, and serves to transfer power from the engine EG to a drive shaft DS of the vehicle.
  • engine internal combustion engine
  • the starting device 1 includes, in addition to the vibration damping device 20 , a front cover 3 coupled to a crankshaft of the engine EG and serving as an input member, a pump impeller (input-side hydraulic transmission element) 4 that is fixed to the front cover 3 and rotates together with the front cover 3 , a turbine runner (output-side hydraulic transmission element) 5 that can rotate coaxially with the pump impeller 4 , a damper hub 7 fixed to an input shaft IS of a transmission (power transmission device) TM, which is an automatic transmission (AT), a continuously variable transmission (CVT), a dual clutch transmission (DCT), a hybrid transmission, or a speed reducer, and serving as an output member, a lockup clutch 8 , a damper device 10 , etc.
  • a transmission power transmission device
  • TM which is an automatic transmission (AT), a continuously variable transmission (CVT), a dual clutch transmission (DCT), a hybrid transmission, or a speed reducer
  • the “axial direction” basically refers to the direction in which the central axis (axis) of the starting device 1 or the damper device 10 (vibration damping device 20 ) extends, unless otherwise specified.
  • the “radial direction” basically refers to the radial direction of the starting device 1 , the damper device 10 , or rotary elements of the damper device 10 etc., namely the direction of a straight line extending from the central axis of the starting device 1 or the damper device 10 perpendicularly to this central axis (in the direction of the radius), unless otherwise specified.
  • the “circumferential direction” basically refers to the circumferential direction of the starting device 1 , the damper device 10 , or the rotary elements of the damper device 10 etc., namely the direction along the rotation direction of the rotary elements, unless otherwise specified.
  • the pump impeller 4 has a pump shell 40 firmly fixed to the front cover 3 , and a plurality of pump blades 41 disposed on the inner surface of the pump shell 40 .
  • the turbine runner 5 has a turbine shell 50 and a plurality of turbine blades 51 disposed on the inner surface of the turbine shell 50 .
  • An inner peripheral part of the turbine shell 50 is fixed to the damper hub 7 via a plurality of rivets.
  • the pump impeller 4 and the turbine runner 5 face each other, and a stator 6 that adjusts the flow of hydraulic oil (working fluid) from the turbine runner 5 to the pump impeller 4 is disposed coaxially between the pump impeller 4 and the turbine runner 5 .
  • the stator 6 has a plurality of stator blades 60 , and the stator 6 is allowed to rotate only in one direction by a one-way clutch 61 .
  • the pump impeller 4 , the turbine runner 5 , and the stator 6 form a torus (annular flow path) in which hydraulic oil is circulated, and function as a torque converter (hydraulic transmission device) having a function to amplify torque.
  • the stator 6 and the one-way clutch 61 may be omitted, and the pump impeller 4 and the turbine runner 5 may function as a fluid coupling.
  • the lockup clutch 8 is a hydraulic multi-plate clutch.
  • the lockup clutch 8 performs lockup coupling, namely couples the front cover 3 to the damper hub 7 , i.e., the input shaft IS of the transmission TM, via the damper device 10 , and releases the lockup coupling.
  • the lockup clutch 8 includes: a lockup piston 80 that is supported by a centerpiece 3 s fixed to the front cover 3 so that the lockup piston 80 can move in the axial direction; a drum portion 11 d that is integral with a drive member 11 , namely an input element of the damper device 10 , and serves as a clutch drum; an annular clutch hub 82 that is fixed to the inner surface of the front cover 3 so as to face the lockup piston 80 ; a plurality of first friction engagement plates (friction plates having a friction material on both surfaces thereof) 83 that are fitted on splines formed in the inner periphery of the drum portion 11 d ; and a plurality of second friction engagement plates (separator plates) 84 that are fitted on splines formed in the outer peripheral surface of the clutch hub 82 .
  • the lockup clutch 8 further includes: an annular flange member (oil chamber defining member) 85 that is attached to the centerpiece 3 s of the front cover 3 so as to be located on the opposite side of the lockup piston 80 from the front cover 3 , namely so as to be located closer to the damper device 10 than the lockup piston 80 is; and a plurality of return springs 86 that are disposed between the front cover 3 and the lockup piston 80 .
  • the lockup piston 80 and the flange member 85 defines an engagement oil chamber 87 , and hydraulic oil (engagement oil pressure) is supplied from a hydraulic control device, not shown, to the engagement oil chamber 87 .
  • the lockup clutch 8 may be a hydraulic single-plate clutch.
  • the damper device 10 includes, as the rotary elements, the drive member (input element) 11 including the drum portion 11 d , an intermediate member (intermediate element) 12 , and a driven member (output element) 15 .
  • the damper device 10 further includes, as torque transmission elements, a plurality of first springs (first elastic bodies) SP 1 and a plurality of second springs (second elastic bodies) SP 2 (e.g., four each in the present embodiment) which are alternately arranged on the same circumference at intervals in the circumferential direction.
  • the first and second springs SP 1 , SP 2 are arc coil springs made of a metal material wound so as to have an axis in the shape of a circular arc when not under load, or straight coil springs made of a metal material wound in a helical pattern so as to have a straight axis when not under load. As shown in the figures, the first and second springs SP 1 , SP 2 may be what is called dual springs.
  • the drive member 11 of the damper device 10 is an annular member including the drum portion 11 d in its outer periphery, and has a plurality of (e.g., four at 90° intervals in the present embodiment) spring contact portions 11 c extended radially inward from its inner peripheral portion at intervals in the circumferential direction.
  • the intermediate member 12 is an annular plate-like member and has a plurality of (e.g., four at 90° intervals in the present embodiment) spring contact portions 12 c extended radially inward from its outer peripheral portion at intervals in the circumferential direction.
  • the intermediate member 12 is rotatably supported by the damper hub 7 , and is disposed radially inside the drive member 11 and surrounded by the drive member 11 .
  • the driven member 15 includes an annular first driven plate 16 and an annular second driven plate 17 coupled to the first driven plate 16 via a plurality of rivets, not shown, so as to rotate therewith.
  • the first driven plate 16 is a plate-like annular member and is disposed closer to the turbine runner 5 than the second driven plate 17 is.
  • the first driven plate 16 together with the turbine shell 50 of the turbine runner 5 is fixed to the damper hub 7 via a plurality of rivets.
  • the second driven plate 17 is a plate-like annular member having a smaller inside diameter than the first driven plate 16 , and an outer peripheral portion of the second driven plate 17 is fastened to the first driven plate 16 via a plurality of rivets, not shown.
  • the first driven plate 16 has: a plurality of (e.g., four in the present embodiment) spring accommodating windows 16 w formed at intervals (regular intervals) in the circumferential direction and each having the shape of a circular arc; a plurality of (e.g., four in the present embodiment) spring support portions 16 a formed at intervals (regular intervals) in the circumferential direction and each extending along the inner peripheral edge of a corresponding one of the spring accommodating windows 16 w ; a plurality of (e.g., four in the present embodiment) spring support portions 16 b formed at intervals (regular intervals) in the circumferential direction and each extending along the outer peripheral edge of a corresponding one of the spring accommodating windows 16 w and facing a corresponding one of the spring support portions 16 a in the radial direction of the first driven plate 16 ; and a plurality of (e.g., four in the present embodiment) spring contact portions 16 c .
  • the plurality of spring contact portions 16 c of the first driven plate 16 are formed so that one spring contact portion 16 c is located between two spring accommodating windows 16 w (spring support portions 16 a , 16 b ) which are adjacent to each other in the circumferential direction.
  • the second driven plate 17 also has: a plurality of (e.g., four in the present embodiment) spring accommodating windows 17 w formed at intervals (regular intervals) in the circumferential direction and each having the shape of a circular arc; a plurality of (e.g., four in the present embodiment) spring support portions 17 a formed at intervals (regular intervals) in the circumferential direction and each extending along the inner peripheral edge of a corresponding one of the spring accommodating windows 17 w ; a plurality of (e.g., four in the present embodiment) spring support portions 17 b formed at intervals (regular intervals) in the circumferential direction and each extending along the outer peripheral edge of a corresponding one of the spring accommodating windows 17 w and facing a corresponding one of the spring support portions 17 a in the radial direction of the second driven plate 17 ; and a plurality of (e.g., four in the present embodiment) spring contact portions 17 c .
  • the plurality of spring contact portions 17 c of the second driven plate 17 are formed so that one spring contact portion 17 c is located between two spring accommodating windows 17 w (spring support portions 17 a , 17 b ) which are adjacent to each other in the circumferential direction.
  • the drive member 11 is rotatably supported by the outer peripheral surface of the second driven plate 17 that is supported by the damper hub 7 via the first driven plate 16 . The drive member 11 is thus aligned with respect to the damper hub 7 .
  • first spring SP 1 and one second spring SP 2 are disposed between two spring contact portions 11 c of the drive member 11 adjacent to each other so that the first and second springs SP 1 , SP 2 are alternately arranged in the circumferential direction of the damper device 10 .
  • Each spring contact portion 12 c of the intermediate member 12 is located between the first and second springs SP 1 , SP 2 that are paired (act in series) as they are disposed between two spring contact portions 11 c adjacent to each other, and contacts the ends of these first and second springs SP 1 , SP 2 .
  • each first spring SP 1 contacts a corresponding one of the spring contact portions 11 c of the drive member 11
  • the other end of each first spring SP 1 contacts a corresponding one of the spring contact portions 12 c of the intermediate member 12
  • one end of each second spring SP 2 contacts a corresponding one of the spring contact portions 12 c of the intermediate member 12
  • the other end of each second spring SP 2 contacts a corresponding one of the spring contact portions 11 c of the drive member 11 .
  • each of the plurality of spring support portions 16 a of the first driven plate 16 supports (guides) the turbine runner 5 sides of a corresponding pair of first and second springs SP 1 , SP 2 from inside in the radial direction.
  • Each of the plurality of spring support portions 16 b supports (guides) the turbine runner 5 sides of a corresponding pair of first and second springs SP 1 , SP 2 from outside in the radial direction.
  • each of the plurality of spring support portions 17 a of the second driven plate 17 supports (guides) the lockup piston 80 sides of a corresponding pair of first and second springs SP 1 , SP 2 from inside in the radial direction.
  • Each of the plurality of spring support portions 17 b supports (guides) the lockup piston 80 sides of a corresponding pair of first and second springs SP 1 , SP 2 from outside in the radial direction.
  • each of the spring contact portions 16 c and the spring contact portions 17 c of the driven member 15 is located between the first and second springs SP 1 , SP 2 that are not paired (do not act in series), and contacts the ends of these first and second springs SP 1 , SP 2 .
  • each first spring SP 1 also contacts a corresponding one of the spring contact portions 16 c and a corresponding one of the spring contact portions 17 c of the driven member 15
  • the other end of each second spring SP 2 also contacts a corresponding one of the spring contact portions 16 c and a corresponding one of the spring contact portions 17 c of the driven member 15
  • the driven member 15 is thus coupled to the drive member 11 via the plurality of first springs SP 1 , the intermediate member 12 , and the plurality of second springs SP 2 , and the first and second springs SP 1 , SP 2 that are paired are coupled in series between the drive member 11 and the driven member 15 via the spring contact portions 12 c of the intermediate member 12 .
  • the distance between the axis of the starting device 1 and the damper device 10 and the axis of each first spring SP 1 is the same as that between the axis of the starting device 1 etc. and the axis of each second spring SP 2 .
  • the damper device 10 of the present embodiment further includes a first stopper that restricts relative rotation between the intermediate member 12 and the driven member 15 and deflection of the second springs SP 2 , and a second stopper that restricts relative rotation between the drive member 11 and the driven member 15 .
  • the first stopper restricts relative rotation between the intermediate member 12 and the driven member 15 when torque that is transferred from the engine EG to the drive member 11 reaches predetermined torque (first threshold value) T 1 smaller than torque T 2 (second threshold value) corresponding to a maximum torsion angle of the damper device 10 .
  • the second stopper restricts relative rotation between the drive member 11 and the driven member 15 when torque that is transferred to the drive member 11 reaches the torque T 2 corresponding to the maximum torsion angle.
  • the damper device 10 thus has two-step (two-stage) damping characteristics.
  • the first stopper may restrict relative rotation between the drive member 11 and the intermediate member 12 and deflection of the first springs SP 1 .
  • the damper device 10 may include a stopper that restricts relative rotation between the drive member 11 and the intermediate member 12 and deflection of the first springs SP 1 , and a stopper that restricts relative rotation between the intermediate member 12 and the driven member 15 and deflection of the second springs SP 2 .
  • the vibration damping device 20 is coupled to the driven member 15 of the damper device 10 and is disposed in a hydraulic transmission chamber 9 filled with hydraulic oil.
  • the vibration damping device 20 includes the first driven plate 16 serving as a support member (first link), a plurality of (e.g., four in the present embodiment) crank members 22 each rotatably coupled to the first driven plate 16 via a first coupling shaft 21 and serving as a restoring force generating member (second link), a single annular inertial mass body (third link) 23 , and a plurality of (e.g., four in the present embodiment) second coupling shafts 24 each coupling a corresponding one of the crank members 22 and the inertial mass body 23 so that the crank member 22 and the inertial mass body 23 can rotate relative to each other.
  • the first driven plate 16 has a plurality of (e.g., four in the present embodiment) protruding support portions 162 formed at intervals (regular intervals) in the circumferential direction so as to protrude radially outward from an outer peripheral surface 161 of the first driven plate 16 .
  • one end of each crank member 22 is rotatably coupled to a corresponding one of the protruding support portions 162 of the first driven plate 16 via the first coupling shaft 21 (see FIG. 3 ).
  • each crank member 22 has two plate members 220 .
  • Each plate member 220 is made of a metal sheet and has a circular arc planar shape.
  • the radius of curvature of the outer peripheral edge of the plate member 220 is designed to be the same as that of the outer peripheral edge of the inertial mass body 23 .
  • the two plate members 220 face each other in the axial direction of the damper device 10 via a corresponding one of the protruding support portions 162 and the inertial mass body 23 and are coupled to each other via the first coupling shaft 21 .
  • the first coupling shaft 21 is a rivet that is inserted through a coupling hole (circular hole) formed in the protruding support portion 162 of the first driven plate 16 and serving as a sliding bearing portion and coupling holes (circular holes) formed in respective plate members 220 and serving as sliding bearing portions and that is clinched at its both ends.
  • the first driven plate 16 (driven member 15 ) and each crank member 22 thus form a turning pair.
  • the first coupling shaft 21 may be inserted through the coupling holes formed in the protruding support portion 162 and one of the two plate members 220 and serving as sliding bearing portions and be supported by (fitted in or fixed to) the other plate member 220 .
  • a rolling bearing such as a ball bearing may be disposed either between the plate member 220 and the first coupling shaft 21 or between the protruding support portion 162 and the first coupling shaft 21 , or both between the plate member 220 and the first coupling shaft 21 and between the protruding support portion 162 and the first coupling shaft 21 .
  • the inertial mass body 23 includes two annular members 230 made of a metal sheet, and the weight of the inertial mass body 23 (two annular members 230 ) is designed to be sufficiently heavier than that of a single crank member 22 .
  • the annular member 230 has a short cylindrical (ring-shaped) body 231 and a plurality of (e.g., four in the present embodiment) protruding portions 232 formed at intervals (regular intervals) in the circumferential direction so as to protrude radially inward from the inner peripheral surface of the body 231 .
  • the two annular members 230 are coupled via a fixing member, not shown, such that the protruding portions 232 face each other in the axial direction of the annular members 230 .
  • Each protruding portion 232 has a guide portion 235 that guides the second coupling shaft 24 coupling the crank member 22 and the inertial mass body 23 .
  • the guide portion 235 is a circular arc-shaped opening and includes a concave guide surface 236 , a convex support surface 237 located closer to the inner periphery of the annular member (first driven plate 16 ) (closer to the center of the annular member 230 ) than the guide surface 236 is and facing the guide surface 236 , and two stopper surfaces 238 located on both sides of the guide surface 236 and the support surface 237 and continuous with the guide surface 236 and the support surface 237 .
  • the guide surface 236 is a concave cylindrical surface with a constant radius of curvature.
  • the support surface 237 is a circular arc-shaped convex surface
  • the stopper surfaces 238 are circular arc-shaped concave surfaces.
  • the guide portion 235 (the guide surface 236 , the support surface 237 , and the stopper surfaces 238 ) is symmetric with respect to a straight line passing through the center of curvature of the guide surface 236 and the center of the annular member 230 (the center of rotation RC of the first driven plate 16 ).
  • a straight line passing through the center of curvature of the guide surface 236 and perpendicular to the protruding portion 232 (annular member 230 ) is defined as an imaginary axis (third coupling shaft) 25 whose position relative to the two annular members 230 , namely the inertial mass body 23 , does not change (which does not move relative to the inertial mass body 23 ).
  • the second coupling shaft 24 is in the shape of a solid (or hollow) round bar and has two projecting portions 24 a that are in the shape of, e.g., a round bar and project axially outward from its both ends. As shown in FIG. 4 , the two projecting portions 24 a of the second coupling shaft 24 are fitted (fixed) in coupling holes (circular holes) formed in the plate members 220 of the crank member 22 .
  • the coupling hole of the plate member 220 in which the projecting portion 24 a is fitted is formed in each plate member 220 so that the central axis of the coupling hole extends coaxially with a straight line passing through the center of gravity G of the crank member 22 (near the central part in the longitudinal direction of the plate member 220 ).
  • the length from the center of the first coupling shaft 21 coupling the first driven plate 16 (protruding support portion 162 ) and the crank member 22 to the center of gravity G of the crank member 22 is thus the same as the axis-to-axis distance (center-to-center distance) between the first coupling shaft 21 and the second coupling shaft 24 coupling the crank member 22 and the inertial mass body 23 .
  • the other end of the crank member 22 (plate members 220 ) is located on the opposite side of the second coupling shaft 24 from the first coupling shaft 21 .
  • Each projecting portion 24 a of the second coupling shaft 24 may be inserted through a coupling hole (circular hole) formed in the plate member 220 of the crank member 22 and serving as a sliding bearing portion. That is, the second coupling shaft 24 may be rotatably supported from both sides by the two plate members 220 , namely the crank member 22 .
  • a rolling bearing such as a ball bearing may be disposed between each plate member 220 and each projecting portion 24 a of the second coupling shaft 24 .
  • the second coupling shaft 24 rotatably supports a cylindrical outer ring 27 via a plurality of rollers (rolling elements) 26 .
  • the outside diameter of the outer ring 27 is designed to be slightly smaller than the interval between the guide surface 236 and the support surface 237 of the guide portion 235 .
  • the second coupling shaft 24 and the outer ring 27 are disposed in corresponding ones of the guide portions 235 of the inertial mass body 23 so as to be supported by the crank member 22 and so that the outer ring 27 rolls on the guide surfaces 236 .
  • the inertial mass body 23 is thus disposed coaxially with the center of rotation RC of the first driven plate 16 so that the inertial mass body 23 can rotate about the center of rotation RC.
  • crank member 22 and the inertial mass body 23 are allowed to rotate relative to each other.
  • Each crank member 22 and the inertial mass body 23 thus form a turning pair.
  • a plurality of balls may be disposed between the second coupling shaft 24 and the outer ring 27 .
  • the first driven plate 16 (driven member 15 ) and each crank member 22 forms a turning pair
  • each crank member 22 and the second coupling shaft 24 guided by the guide portions 235 of the inertial mass body 23 forms a turning pair.
  • the inertial mass body 23 is rotatably disposed about the center of rotation RC of the first driven plate 16 .
  • each second coupling shaft 24 is guided by the guide portions 235 of the inertial mass body 23 and moves with motion of the second link to swing (reciprocating rotary motion) about the first coupling shaft 21 while maintaining a constant axis-to-axis distance to the first coupling shaft 21 and to swing (reciprocating rotary motion) about the imaginary axis 25 while maintaining a constant axis-to-axis distance to the imaginary axis 25 .
  • each crank member 22 swings about the first coupling shaft 21 with the movement of the second coupling shaft 24 , and the imaginary axis 25 and the inertial mass body 23 swing about the moving second coupling shaft 24 and swing (reciprocating rotary motion) about the center of rotation RC of the first driven plate 16 .
  • the first driven plate 16 , the crank member 22 , the inertial mass body 23 , the first and second coupling shafts 21 , 24 , and the guide portions 235 thus substantially form a four-bar linkage mechanism in which the first driven plate 16 serves as a fixed link.
  • the first driven plate 16 , the crank members 22 , the inertial mass body 23 , the second coupling shafts 24 , and the guide portions 235 of the inertial mass body 23 are configured to satisfy the relationship L1+L2>L3+L4, where “L1” represents the axis-to-axis distance between the center of rotation RC of the first driven plate 16 and the first coupling shaft 21 , “L2” represents the axis-to-axis distance between the first coupling shaft 21 and the second coupling shaft 24 , “L3” represents the axis-to-axis distance between the second coupling shaft 24 and the imaginary axis 25 , and “L4” represents the axis-to-axis distance the imaginary axis 25 and the center of rotation RC (see FIG.
  • the axis-to-axis distance L3 between the second coupling shaft 24 and the imaginary axis 25 (the radius of curvature of the guide surface 236 minus the radius of the outer ring 27 ) is designed to be shorter than the axis-to-axis distances L1, L2, and L4 and is set as short as possible within the range that does not affect the operation of each crank member 22 and the inertial mass body 23 .
  • the first driven plate 16 (protruding support portion 162 ) serving as a first link is configured so that the axis-to-axis distance L1 between the center of rotation RC and the first coupling shaft 21 is longer than the axis-to-axis distances L2, L3, and L4.
  • the vibration damping device 20 of the present embodiment satisfies the relationship L1>L4>L2>L3, and the first driven plate 16 , the crank member 22 , the inertial mass body 23 , the first and second coupling shafts 21 , 24 , and the guide portions 235 substantially form a double lever mechanism in which the first driven plate 16 facing a line segment (imaginary link) connecting the second coupling shaft 24 and the imaginary axis 25 serves as a fixed link.
  • An “equilibrium (balanced) state” of the vibration damping device 20 is the state where the resultant force of the sum of centrifugal forces acting on the components of the vibration damping device 20 and the forces acting on the centers of the first and second coupling shafts 21 , 24 of the vibration damping device 20 and the center of rotation RC is zero. As shown in FIG. 2 , when the vibration damping device 20 is in the equilibrium state, the center of the second coupling shaft 24 , the center of the imaginary axis 25 , and the center of rotation RC of the first driven plate 16 are located on a straight line.
  • the vibration damping device 20 of the present embodiment is also configured to satisfy 60° ⁇ 120°, more preferably 70° ⁇ 90°, where “ ⁇ ” represents the angle formed by the direction from the center of the first coupling shaft 21 toward the center of the second coupling shaft 24 and the direction from the center of the second coupling shaft 24 toward the center of rotation RC when the vibration damping device 20 is in the equilibrium state where the center of the second coupling shaft 24 , the center of the imaginary axis 25 , and the center of rotation RC of the first driven plate 16 are located on a straight line.
  • the first driven plate 16 (driven member 15 ) of the damper device 10 also rotates in the same direction as the front cover 3 about the axis of the starting device 1 .
  • each crank member 22 and the inertial mass body 23 of the vibration damping device 20 swing with respect to the first driven plate 16 as shown in FIGS. 5A, 5B, and 5C .
  • Vibration in opposite phase from vibration transmitted from the engine EG to the drive member 11 is thus applied from the swinging inertial mass body 23 to the first driven plate 16 via each guide portion 235 , each second coupling shaft 24 , and each crank member 22 , whereby vibration of the first driven plate 16 can be dampened. That is, the vibration damping device 20 is configured to have an order according to the order of vibration that is transmitted from the engine EG to the first driven plate 16 (vibration order: 1.5th order in the case where the engine EG is, e.g., a three-cylinder engine, and second order in the case where engine EG is, e.g., a four-cylinder engine).
  • the vibration damping device 20 damps vibration that is transmitted from the engine EG to the first driven plate 16 , regardless of the rotational speed of the engine EG (first driven plate 16 ). Vibration can thus be very satisfactorily damped by both the damper device 10 and the vibration damping device 20 while an increase in weight of the damper device 10 is retained.
  • a four-bar linkage mechanism can be configured without using a link coupled to both the crank member 22 and the inertial mass body 23 , namely a connecting rod of a common four-bar linkage mechanism.
  • An increase in overall weight and size of the device is thus satisfactorily restrained.
  • a restoring force acting on the crank member 22 is restrained from being reduced due to the center of gravity G of the crank member 22 being moved toward the center of rotation RC by an increase in weight (moment of inertia) of the connecting rod. Satisfactory vibration damping capability is thus achieved.
  • the axis-to-axis distance L 3 between the second coupling shaft 24 and the imaginary axis 25 can be set with greater flexibility, whereby the axis-to-axis distance L3 can be easily reduced.
  • the vibration damping capability of the vibration damping device 20 can thus be easily improved by adjusting the axis-to-axis distance L3.
  • the vibration damping capability of the vibration damping device 20 is thus improved while an increase in weight of the crank member 22 is restrained. As a result, in the vibration damping device 20 , the vibration damping capability is further improved while restraining an increase in overall weight and size of the device.
  • this vibration damping device can be said to correspond to a centrifugal pendulum vibration absorbing device.
  • centrifugal pendulum vibration absorbing devices the deflection angle of a pendulum mass body increases with an increase in amplitude of vibration of input torque that is transferred to a support member of the pendulum mass body. As the deflection angle increases, a restoring force that tries to return the pendulum mass body to the equilibrium state (balanced position) decreases accordingly.
  • the amount of decrease in restoring force namely in equivalent stiffness of the centrifugal pendulum vibration absorbing device
  • an effective order which is an order of vibration that can be most satisfactorily damped by the centrifugal pendulum vibration absorbing device
  • the larger the amount of decrease in effective order (the difference from the excitation order) is, the more the vibration damping capability is degraded.
  • the centrifugal pendulum vibration absorbing devices are therefore typically designed to minimize the decrease in effective order which occurs as the deflection angle increases.
  • the vibration damping device 20 when the amplitude ⁇ of vibration of torque that is transferred from the drive member 11 to the driven member 15 (hereinafter referred to as the “input torque”) increases and the deflection angle of the inertial mass body 23 increases accordingly, deviation is produced between the order of vibration that is supposed to be damped by the vibration damping device 20 , which is the excitation order q tag of the engine EG, and the effective order q eff , which is the order of vibration that is most satisfactorily damped by the vibration damping device 20 .
  • the effective order q eff may become smaller than the excitation order q tag of the engine EG or larger than the excitation order q tag of the engine EG depending on the specifications of the vibration damping device.
  • the inventors therefore first carried out simulations to find such combinations of the axis-to-axis distances L2, L3, L4 and the length Lg (length from the center of the first coupling shaft 21 to the center of gravity G of the crank member 22 ) that do not vary the effective order q eff even if the amplitude ⁇ of vibration of input torque varies.
  • the mass m of the crank members 22 , the moment of inertia (inertia) J of the inertial mass body 23 , the number of cylinders n of the engine EG, and the axis-to-axis distance L1 that depends on the requirements for mounting the vibration damping device 20 , etc. were the same in the simulations.
  • a plurality of models of the vibration damping device 20 with various axis-to-axis distances L2, L3, L4 and various lengths Lg were used in the simulations.
  • the initial state was the state where the inertial mass body 23 had been rotated by a certain initial angle (angle corresponding to the deflection angle of the inertial mass body 23 about the center of rotation RC) about the center of rotation RC from its position in the equilibrium state.
  • torque containing no vibration component was applied to the first driven plate 16 at every plurality of initial angles to rotate the first driven plate 16 at a constant rotational speed (e.g., 1000 rpm) so as to cause the inertial mass body 23 etc. to swing at a frequency corresponding to the initial angle.
  • the results of the simulations show that, in the case where the vibration damping device 20 satisfies the relationship as given by the following expression (1), the effective order q eff is kept approximately constant even if the amplitude ⁇ of vibration of input torque varies.
  • “ ⁇ ,” “ ⁇ ,” and “ ⁇ ” are constants determined by the simulations.
  • the results of analysis carried out by the inventors show that, in the case where the vibration damping device 20 satisfies the relationship as given by the following expression (2), the effective order q eff increases as the amplitude ⁇ of vibration of input torque increases, and in the case where the vibration damping device 20 satisfies the relationship as given by the following expression (3), the effective order q eff decreases as the amplitude ⁇ of vibration of input torque increases.
  • the results of the analysis show that, in the vibration damping device 20 that satisfies any one of the expressions (1), (2), and (3), the value to which the effective order q eff converges (hereinafter referred to as the “reference order q ref ”) as the amplitude ⁇ of vibration of input torque decreases is varied by changing the mass m of the crank members 22 and the moment of inertia J of the inertial mass body 23 .
  • L 4/( L 3+ L 4) ⁇ ( Lg/L 2)+ ⁇ n+ ⁇ (1)
  • the inventors also considered the relationship between the reference order q ref and the vibration damping capability of the vibration damping device 20 based on the results of the simulations and analysis described above.
  • the inventors evaluated the relationship between the rotational speed Ne of the engine EG (a three-cylinder engine in this example) and the torque fluctuation T Fluc of a final object whose vibration is to be damped (the drive shaft DS in this example) by numerical analysis using LMS Imagine.Lab Amesim (registered trademark) from Simens AG.
  • the amount of deviation of the effective order q eff from the excitation order q tag is a value obtained by subtracting the excitation order q tag from the effective order q eff at the time the amplitude ⁇ of vibration of input torque reaches its maximum value and the deflection angle of the inertial mass body 23 reaches its maximum value.
  • the dynamic model used for the analysis is a dynamic model simulating the vehicle structure (see FIG. 1 ) from the engine EG to wheels W which includes the vibration damping device 20 and the final object whose vibration is to be damped (the drive shaft DS in this example).
  • This dynamic model is a multi-degree-of-freedom dynamic model produced in view of non-linearity of objects whose vibration is to be damped and the vibration damping device 20 .
  • the moment of inertia J 1 of the driven member 15 , the moment of inertia J 2 of the inertial mass body 23 , the axis-to-axis distances L3, L4, and the mass m of all the crank members 22 out of the design parameters of the dynamic model were varied as appropriate in view of mountability of the vibration damping device 20 on the starting device 1 (vehicle) etc. so that the plurality of models of the vibration damping device 20 had different reference orders q ref from each other. That is, in the analysis, the design parameters of the dynamic model other than the moments of inertia J 1 , J 2 and the axis-to-axis distances L3, L4, and the mass m had fixed values.
  • FIG. 6 shows the analysis results of the relationship between the rotational speed Ne and the torque fluctuation T Fluc of the drive shaft DS in a plurality of models M 0 to M 11 of the vibration damping device 20 having different reference orders q ref from each other.
  • This figure shows the analysis results of the torque fluctuation T Fluc (vibration level) of the drive shaft DS with torque being transferred from the engine EG to the driven member 15 by the lockup coupling.
  • the models M 1 to M 7 are models for which the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m were determined so that the reference order q ref is increased in increments of 0.005.
  • the reference order q ref of the model M 7 1.535.
  • the reference order q ref was varied among the models M 0 to M 11 by using the same moment of inertia J 1 of the driven member 15 , the same axis-to-axis distances L3, L4, and the same mass m of the crank members 22 in all the models M 0 to M 11 and varying the moment of inertia J 2 of the inertial mass body 23 among the models M 0 to M 11 .
  • the reference order q ref may be adjusted by using the same moments of inertia J 1 , J 2 and the same axis-to-axis distances L3, L4 and varying the mass m of the crank members 22 .
  • the torque fluctuation T Fluc of the drive shaft DS at around a lockup rotational speed Nlup of the lockup clutch Bis larger than an allowable value shown by dashed line in the figure, and a peak value (see circles in the figure) of the torque fluctuation T Fluc in a predetermined rotational speed range (e.g., 1000 to 2000 rpm) in a lockup region is also larger than the allowable value.
  • a predetermined rotational speed range e.g. 1000 to 2000 rpm
  • the torque fluctuation T Fluc of the drive shaft DS at around the lockup rotational speed Nlup is sufficiently small and is smaller than both the allowable value and the torque fluctuation T Fluc of the drive shaft DS at around the lockup rotational speed Nlup in the model M 0 in which the rate ⁇ is 0%, and a peak value of the torque fluctuation T Fluc in the predetermined rotational speed range in the lockup region is also sufficiently small.
  • the vibration damping capability can be more improved by making the reference order q ref , namely the value to which the effective order q eff converges as the amplitude ⁇ of vibration of input torque decreases, larger than the excitation order q tag of the engine EG than by making the reference order q ref the same as the excitation order q tag .
  • the vibration damping device 20 of the present embodiment is designed so that the reference order q ref is larger than the excitation order q tag of the engine EG.
  • the vibration damping device is designed to satisfy 1.00 ⁇ q tag ⁇ q ref ⁇ 1.03 ⁇ q tag , more preferably 1.01 ⁇ q tag ⁇ q ref ⁇ 1.02 ⁇ q tag .
  • the peak value of the torque fluctuation T Fluc in the predetermined rotational speed range in the lockup region is thus made closer to a value (target value or target range) around the minimal value MIN, whereby the vibration damping capability of the vibration damping device 20 including the crank members 22 and the inertial mass body 23 that swings with motion of the crank members 22 can be very satisfactorily improved.
  • the vibration damping device 20 is designed by a procedure as shown in FIG. 8 . That is, when designing the vibration damping device 20 , a dynamic model as described above, namely, a dynamic model simulating the vehicle structure from the engine EG to the wheels W which includes the vibration damping device 20 and the final object whose vibration is to be damped, is first created (step S 100 ).
  • the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m are set as design parameters that make the reference order q ref the same as or slightly larger than the excitation order q tag (step S 110 ).
  • numerical calculation (calculation of the response described above) is performed by software such as LMS Imagine.Lab Amesim (step S 120 ) and torque fluctuation T Fluc in the drive shaft DS whose vibration is to be damped is derived (S 130 ).
  • step S 130 it is determined whether a minimal value MIN of a peak value of the torque fluctuation T Fluc in a predetermined rotational speed range (e.g., 1000 to 2000 rpm) can be recognized or not, namely whether the peak value of the torque fluctuation T Fluc has switched from decreasing to increasing or not (step S 140 ).
  • a minimal value MIN of a peak value of the torque fluctuation T Fluc in a predetermined rotational speed range e.g., 1000 to 2000 rpm
  • step S 140 If it is determined in step S 140 that the minimal value MIN of the peak value of the torque fluctuation T Fluc cannot be recognized, the routine returns to step S 110 , where the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters are reset (changed) so that the reference order q ref becomes larger than the reference order q ref determined by the design parameters set in the previous step S 110 . At this time, as described above, only the moment of inertia J 2 or the mass m may be changed in order to increase the reference order q ref . After the design parameters are reset in this manner, the processing of the above steps S 120 to S 140 is performed again.
  • step S 110 the design parameters are set (changed) in step S 110 so that the reference order q ref increases, and the processing of steps S 120 and S 130 is repeated until it is determined in step S 140 that the minimal value MIN of the peak value of the torque fluctuation T Fluc can be recognized.
  • step S 140 If it is determined in step S 140 that the minimal value MIN of the peak value of the torque fluctuation T Fluc can be recognized, the design parameters that make the peak value of the torque fluctuation T Fluc equal to a target value, namely the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters that make the peak value closer to the minimal value MIN or make the peak value equal to the minimal value MIN, are selected (step S 150 ).
  • step S 150 the design parameters (e.g., the moment of inertia J 2 or the mass m) are changed so that the reference order q ref becomes larger than the reference order q ref determined by the design parameters immediately before the peak value of the torque fluctuation T Fluc starts to increase, and the torque fluctuation T Fluc is derived in a manner similar to that of steps S 120 and S 130 to find the design parameters that make the peak value of the torque fluctuation T Fluc in the predetermined rotational speed range equal to the minimal value MIN or to a value close to the minimal value MIN.
  • the design parameters e.g., the moment of inertia J 2 or the mass m
  • the values of the design parameters thus selected namely the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m, are determined as design values (step S 160 ).
  • the design for making the reference order q ref larger than the excitation order q tag and making the peak value of the torque fluctuation T Fluc closer to the minimal value MIN is thus completed.
  • the reference order q ref of the vibration damping device 20 is determined according to the values of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m determined as design values.
  • the design parameters may be automatically set (changed) by software, the design parameters that have been set separately may be input to software, or a plurality of sets of design parameters (J 1 , J 2 , L3, L4, m) determined in advance so that the reference order q ref increases may be sequentially set as parameters for numerical analysis.
  • the design parameters that makes the peak value of the torque fluctuation T Fluc closer to the minimal value MIN may be determined from the design parameters immediately before the peak value of the torque fluctuation T Fluc starts to increase and the design parameters at the time the peak value is starting to increase, in view of mountability of the vibration damping device 20 on the starting device 1 (vehicle) etc.
  • a plurality of design parameters that cause the peak value of the torque fluctuation T Fluc in the predetermined rotational speed range which is derived in step S 130 of FIG. 8 to be included in a predetermined target value (allowable range, near the minimal value MIN in FIG. 7 ) may be extracted and a parameter that minimizes the peak value of the torque fluctuation T Fluc may be selected as a design value from the plurality of design parameters.
  • the target value is set to a value smaller than the torque fluctuation T Fluc of an object whose vibration is to be damped (drive shaft DS) at the time the reference order q ref is equal to the excitation order q tag of the engine EG.
  • step S 100 of FIG. 8 the moment of inertias J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters may be set according to the above expression (2) so that the effective order q eff gradually increases as the amplitude ⁇ of vibration of input torque that is transferred from the engine EG to the driven member 15 increases.
  • the vibration damping device 20 being designed according to the above expression (2) so that the effective order q eff increases as the amplitude ⁇ of vibration of input torque that is transferred from the engine EG to the driven member 15 increases, the rotational speed range in which the vibration damping capability is degraded due to any deviation of the effective order q eff which occurs as the amplitude ⁇ increases can be shifted toward higher rotation.
  • This restrains degradation in vibration damping capability due to the deviation of the effective order q eff from occurring in the range where the rotational speed Ne is relatively low in the lockup region, and further improves the vibration damping capability in the range where the rotational speed Ne of the engine EG is low.
  • the vibration damping device 20 may be designed so that the effective order q eff does not vary even if the amplitude ⁇ of vibration of input torque that is transferred from the engine EG to the driven member 15 varies (so that the rate ⁇ is equal to 0%). This also restrains the vibration damping capability from being degraded by the deviation of the effective order q eff in the range where the rotational speed Ne is relatively low in the lockup region, and improves the vibration damping capability in the range where the rotational speed Ne of the engine EG is low.
  • the inventors formed an equation of motion as given by the following expression (4) for the driven member 15 that is one of the objects whose vibration is to be damped, and also formed an equation of motion as given by the following expression (5) for the entire vibration damping device 20 .
  • J 1 represents the moment of inertia of the driven member 15 (first and second driven plates 16 , 17 )
  • J 2 represents the moment of inertia of the inertial mass body 23
  • m represents the mass (sum) of all (four in the present embodiment) of the crank members 22
  • k represents stiffness (spring constant) of a system from the engine EG to the driven member 15 .
  • L 3 represents the distance between the center of gravity G of the crank member 22 and the fulcrum of swinging of the crank member 22 with respect to the inertial mass body 23 , namely the imaginary axis 25
  • L 4 represents the distance between the fulcrum of swinging of the crank member 22 with respect to the inertial mass body 23 , namely the imaginary axis 25
  • the center of rotation RC represents the distance between the fulcrum of swinging of the crank member 22 with respect to the inertial mass body 23 , namely the imaginary axis 25 , and the center of rotation RC.
  • the distance L 3 is the same as the axis-to-axis distance L3
  • the distance L 4 is the same as the axis-to-axis distance L4.
  • ⁇ represents the angular velocity of steady rotation of the driven member 15
  • ⁇ represents angular displacement of the driven member 15 from the steady rotation state
  • ⁇ represents the angle formed by a reference line defined on the first driven plate 16 (driven member 15 ) and extending in the radial direction of the first driven plate 16 through the center of rotation RC and a line segment from the fulcrum of swinging of the crank member 22 with respect to the inertial mass body 23 , namely the imaginary axis 25 , to the center of gravity G of the crank member 22 (in the present embodiment, the center of the second coupling shaft 24 ) (see FIG.
  • “ ⁇ ” represents the angle formed by the reference line and a line segment from the center of rotation RC to the fulcrum of swinging of the crank member 22 with respect to the inertial mass body 23 , namely the imaginary axis 25 (see FIG. 5A ).
  • “Td” represents the torque that is transferred from the engine EG to the driven member 15
  • “ ⁇ ” represents the frequency of vibration that is transferred from the engine EG to the driven member 15 .
  • the symbol “′” represents the first derivative with respect to the angle ⁇
  • the symbol “′′” represents the second derivative with respect to the angle ⁇ .
  • the reference line is a straight line (L1) passing through the center of rotation RC and the center of the first coupling shaft 21 .
  • the angle ⁇ is therefore an angle formed by the straight line (L1) passing through the center of rotation RC and the center of the first coupling shaft 21 and a line segment (L3) from the center of the second coupling shaft 24 to the imaginary axis 25
  • the angle ⁇ is an angle formed by the straight line (L1) passing through the center of rotation RC and the center of the first coupling shaft 21 and a line segment (L4) from the center of rotation RC to the imaginary axis 25 .
  • the angle ⁇ see FIG.
  • the angular displacement ⁇ of the driven member 15 from the steady rotation state can be regarded as zero.
  • the natural angular frequency ⁇ a of the vibration damping device 20 near the equilibrium state as given by the following expression (8) can be obtained by ignoring the first term on the left side of the expression (7).
  • the reference order q ref namely the value to which the effective order q eff converges as the amplitude ⁇ of vibration of input torque decreases, as given by the following expression (9) can be obtained from the expression (8).
  • the angular displacement ⁇ of the driven member 15 from the steady rotation state, the angle formed by the reference line and the line segment (L3) from the center of gravity of the crank member 22 (the center of the second coupling shaft 24 ) to the imaginary axis 25 , and the angle ⁇ formed by the reference line and the line segment (L4) from the center of rotation RC to the imaginary axis 25 at the time certain torque Td is transferred from the engine EG to the driven member 15 by the lockup coupling can be derived for a plurality of engine speeds Ne.
  • the angle ⁇ formed by the reference line and the line segment from the center of the first coupling shaft 21 to the center of the second coupling shaft 24 can be calculated for the plurality of engine speeds Ne based on the derived angle ⁇ etc.
  • torque fluctuation that is transmitted to a final object whose vibration is to be damped (the drive shaft DS in this example) can be derived for the plurality of engine speeds Ne based on the angular displacement ⁇ and the angles ⁇ , ⁇ , and ⁇ obtained from the equations of motion as given by the expressions (4) and (5).
  • the vibration damping device 20 can be designed so that the reference order q ref is larger than the excitation order q tag of the engine EG.
  • the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m are first set as design parameters that make the reference order q ref equal to the excitation order q tag or slightly larger than the excitation order q tag (step S 200 ).
  • step S 200 the equations of motion as given by the expressions (4) and (5) are solved for the design parameters set in step S 200 to derive the angular displacement ⁇ and the angles ⁇ and ⁇ (and the angle ⁇ ) at the time certain torque Td is transferred to the driven member 15 for a plurality of engine speeds Ne (step S 210 ).
  • Numerical calculation using, e.g., a dynamic model of a system from the driven member 15 to the wheels W in a vehicle having the starting device 1 mounted thereon is then performed to derive torque fluctuation that is transmitted to an object whose vibration is to be damped (the drive shaft DS in this example) for the plurality of engine speeds Ne based on the angular displacement ⁇ and the angles ⁇ , ⁇ , and ⁇ obtained from the equations of motion as given by the expressions (4) and (5) (step S 220 ).
  • step S 220 it is determined whether a minimal value MIN of a peak value of torque fluctuation T Fluc in a predetermined rotational speed range (e.g., 1000 to 2000 rpm) can be recognized or not, namely whether the peak value of the torque fluctuation T Fluc has switched from decreasing to increasing or not, in a manner similar to that of step S 140 of FIG. 8 (step S 230 ).
  • a predetermined rotational speed range e.g. 1000 to 2000 rpm
  • step S 230 If it is determined in step S 230 that the minimal value MIN of the peak value of the torque fluctuation T Fluc cannot be recognized, the routine returns to step S 200 , where the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters are reset (changed) so that the reference order q ref becomes larger than the reference order q ref determined by the design parameters set in the previous step S 200 . At this time, as described above, only the moment of inertia J 2 or the mass m may be changed in order to increase the reference order q ref . After the design parameters are reset in this manner, the processing of the above steps S 210 to S 230 is performed again.
  • step S 200 the design parameters are set (changed) in step S 200 so that the reference order q ref increases, and the processing of steps S 210 and S 220 is repeated until it is determined in step S 230 that the minimal value MIN of the peak value of the torque fluctuation T Fluc can be recognized.
  • step S 230 If it is determined in step S 230 that the minimal value MIN of the peak value of the torque fluctuation T Fluc can be recognized, the design parameters that make the peak value of the torque fluctuation T Fluc equal to a target value, namely the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters that make the peak value closer to the minimal value MIN or make the peak value equal to the minimal value MIN, are selected in a manner similar to that of step S 150 of FIG. 8 (step S 240 ).
  • step S 240 the design parameters (e.g., the moment of inertia J 2 or the mass m) are changed so that the reference order q ref becomes larger than the reference order q ref determined by the design parameters immediately before the peak value of the torque fluctuation T Fluc starts to increase, and the torque fluctuation T Fluc is derived in a manner similar to that of steps S 210 and S 220 to find the design parameters that make the peak value of the torque fluctuation T Fluc in the predetermined rotational speed range equal to the minimal value MIN or to a value close to the minimal value MIN.
  • the design parameters e.g., the moment of inertia J 2 or the mass m
  • the values of the design parameters thus selected namely the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m, are determined as design values (step S 250 ).
  • the design for making the reference order q ref larger than the excitation order q tag and making the peak value of the torque fluctuation T Fluc closer to the minimal value MIN is thus completed.
  • the reference order q ref of the vibration damping device 20 is determined according to the values of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m determined as design values.
  • the design parameters may be automatically set (changed) by software, the design parameters that have been set separately may be input to software, or a plurality of sets of design parameters (J 1 , J 2 , L3, L4, m) determined in advance so that the reference order q ref increases may be sequentially set as parameters to be used for numerical analysis.
  • the design parameters that makes the peak value of the torque fluctuation T Fluc closer to the minimal value MIN may be determined from the design parameters immediately before the peak value of the torque variation T Fluc starts to increase and the design parameters at the time the peak value is starting to increase, in view of mountability of the vibration damping device 20 on the starting device 1 (vehicle) etc.
  • a plurality of design parameters that cause the peak value of the torque fluctuation T Fluc in the predetermined rotational speed range which is derived in step S 220 of FIG. 9 to be included in a predetermined target value may be extracted and a parameter that minimizes the peak value of the torque fluctuation T Fluc may be selected as a design value from the plurality of design parameters.
  • the target value is set to a value smaller than the torque fluctuation T Fluc of an object whose vibration is to be damped (drive shaft DS) at the time the reference order q ref is equal to the excitation order q tag of the engine EG.
  • step S 200 of FIG. 9 the moment of inertias J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters may be set according to the above expression (2) so that the effective order q eff gradually increases as the amplitude ⁇ of vibration of input torque that is transferred from the engine EG to the driven member 15 increases. This restrains the vibration damping capability from being degraded by the deviation of the effective order q eff in the range where the rotational speed Ne is relatively low in the lockup region, and further improves the vibration damping capability in the range where the rotational speed Ne of the engine EG is low.
  • the reference order q ref can be easily made to fall within a proper range that is larger than the excitation order q tag of the engine EG, and the vibration damping capability of the vibration damping device 20 can be improved.
  • the reference order q ref that is determined by the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m applied to the numerical calculation and by the expression (9) is larger than the excitation order q tag of the engine EG.
  • the expression (9) therefore need not necessarily be used to determine a plurality of combinations of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m.
  • torque fluctuation of the rotary element may be derived from the angular displacement ⁇ and the angles ⁇ and ⁇ which are obtained by solving the equations of motion as given by the above expressions (4) and (5).
  • the equation (9) may be used to set a plurality of combinations of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m as design parameters in step S 110 of FIG. 8 .
  • the vibration damping device 20 may be configured to satisfy the relationship Lg>L2.
  • the center of gravity G need not necessarily be located on a straight line passing through the centers of the first and second coupling shafts 21 , 24 .
  • the guide portions 235 may be formed in the crank members 22 , and the second coupling shafts 24 may be supported by the inertial mass body 23 .
  • the guide portion 235 includes the convex support surface 237 facing the guide surface 236 and the stopper surfaces 238 . However, the support surface 237 and the stopper surfaces 238 may be omitted as shown in FIG. 11 .
  • a guide portion 235 V formed in a protruding portion 232 of an annular member 230 V shown in FIG. 11 is an approximately semicircular cutout having a concave (concave cylindrical) guide surface 236 with a constant radius of curvature. This configuration simplifies the structure of the guide portion 235 V that guides the second coupling shaft 24 and thus the structure of the vibration damping device 20 .
  • a guide portion similar to the guide portion 235 V may be formed in the plate members 220 of the crank member 22 .
  • the guide surface 236 may be, e.g., a concave surface whose radius of curvature varies stepwise or gradually as long as the guide surface 236 moves the second coupling shaft 24 in a manner described above.
  • the annular inertial mass body 23 may be configured to be rotatably supported (aligned) by the first driven plate 16 . This configuration allows the inertial mass body 23 to smoothly swing about the center of rotation RC of the first driven plate 16 when the crank members 22 swing.
  • the annular inertial mass body 23 may be replaced with a plurality of (e.g., four) mass bodies having the same specifications (dimensions, weight, etc.).
  • the mass bodies may be, e.g., metal sheets having a circular arc planar shape and each coupled to the first driven plate 16 via the crank member 22 (two plate members 220 ), the second coupling shaft 24 , and the guide portions 235 so that the mass bodies are located at intervals (regular intervals) in the circumferential direction in the equilibrium state and swing about the center of rotation RC.
  • guide portions each receiving a centrifugal force (centrifugal oil pressure) acting on the mass body and guiding the mass body so that the mass body swings about the center of rotation RC may be formed in the outer peripheral portion of the first driven plate 16 .
  • the vibration damping device 20 may include dedicated support members (first links) each swingably supporting the crank member 22 and forming a turning pair with the crank member 22 and also forming a turning pair with the inertial mass body 23 . That is, the crank members 22 may be indirectly coupled to a rotary element via the dedicated support members as first links. In this case, the support members of the vibration damping device 20 need only to be coaxially coupled to a rotary element whose vibration is to be damped, such as, e.g., the drive member 11 , the intermediate member 12 , or the first driven plate 16 of the damper device 10 , so that the support members rotate together with the rotary element.
  • the vibration damping device 20 with this configuration can satisfactorily damp vibration of the rotary element.
  • each connecting rod 35 is rotatably coupled to the crank member 22 via a second coupling shaft 24 X and is rotatably coupled to a protruding portion 232 of an inertial mass body 23 X via a third coupling shaft 30 .
  • This vibration damping device 20 X also has functions and effects similar to those of the vibration damping device 20 when designed based on the above expression (1) or (2).
  • the design values of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m can be determined by, e.g., numerical calculation using software such as LMS Imagine.Lab Amesim described above or numerical calculation using the equations of motion as given by the above expressions (4) and (5) etc.
  • FIG. 13 is an enlarged view of a further vibration damping device 20 Y of the present disclosure
  • FIG. 14 is an enlarged sectional view of a main part of the vibration damping device 20 Y.
  • the vibration damping device 20 Y shown in these figures includes a driven plate 16 Y configured similarly to the first driven plate 16 and serving as a support member, a plurality of (e.g., four in the present embodiment) weight bodies 22 Y each rotatably coupled to the first driven plate 16 via a coupling shaft (coupling member) 214 and serving as a restoring force generating member, and a single annular inertial mass body 23 Y coupled to the driven plate 16 Y and the weight bodies 22 Y via the coupling shafts 214 .
  • the driven plate 16 Y has a plurality of (e.g., four at 90° intervals in the present embodiment) long holes (through holes) 16 h (first guide portions) formed in its outer peripheral portion at intervals (regular intervals) in the circumferential direction.
  • each long hole 16 h guides the coupling shaft 214 formed in the shape of a solid (or hollow) round bar, namely guides the weight body 22 Y, and is formed in the driven plate 16 Y so that the central axis of the long hole 16 h in the longitudinal direction extends in the radial direction of the driven plate 16 Y and passes through the center of rotation RC.
  • each weight body 22 Y has two plate members 220 Y coupled to each other via the coupling shaft 214 .
  • each plate member 220 Y is made of a metal sheet and has the shape of a disc.
  • the coupling shaft 214 is fixed (coupled) to the two plate members 220 Y so that its axis passes through the center of gravity G of the weight body 22 Y.
  • the guide portion 235 Y includes a concave guide surface 236 , a convex support surface 237 located closer to the inner periphery of the annular member 230 Y (closer to the center of the annular member 230 Y) than the guide surface 236 is and facing the guide surface 236 , and two stopper surfaces 238 located on both sides of the guide surface 236 and the support surface 237 and continuous with the guide surface 236 and the support surface 237 .
  • the guide surface 236 is a concave cylindrical surface with a constant radius of curvature.
  • the support surface 237 is a circular arc-shaped convex surface
  • the stopper surfaces 238 are circular arc-shaped concave surfaces.
  • the interval between the guide surface 236 and the support surface 237 is designed to be slightly larger than the outside diameter of the coupling shaft 214 .
  • the guide portion 235 Y (the guide surface 236 , the support surface 237 , and the stopper surfaces 238 ) is symmetric with respect to a straight line passing through the center of curvature of the guide surface 236 and the center of the annular member 230 Y (the center of rotation RC of the driven plate 16 Y).
  • the two annular members 230 Y are disposed on both sides in the axial direction of the driven plate 16 Y, one on each side, and coaxially with the driven plate 16 Y so that corresponding ones of the guide portions 235 Y face each other in the axial direction of the annular members 230 Y.
  • the inner peripheral surfaces of the two annular members 230 Y are supported by a plurality of projections 16 p (see FIG. 13 ) formed on the driven plate 16 Y so as to project in the axial direction.
  • Each annular member 230 Y (inertial mass body 23 Y) is thus supported by the driven plate 16 Y so that it can rotate about the center of rotation RC.
  • the two plate members 220 Y are disposed so as to face each other in the axial direction via the corresponding driven plate 16 Y and the two annular members 230 , and are coupled to each other via the coupling shafts 214 .
  • each coupling shaft 214 coupling the two plate members 220 Y extends through a corresponding one of the long holes 16 h in the driven plate 16 Y and a corresponding one of the guide portions 235 Y in each of the two annular members 230 Y.
  • each coupling shaft 214 can move along both a corresponding one of the long holes 16 h in the driven plate 16 Y and corresponding ones of the guide portions 235 Y in the inertial mass body 23 Y.
  • each weight body 22 Y (coupling shaft 214 ) forms a sliding pair with the driven plate 16 Y and the inertial mass body 23 Y, and the driven plate 16 Y and the inertial mass body 23 Y form a turning pair.
  • the driven plate 16 Y having the long holes 16 h , the plurality of weight bodies 22 Y, and the inertial mass body 23 Y having the guide portions 235 Y thus form a slider crank mechanism (double slider crank chain).
  • the equilibrium state of the vibration damping device 20 Y is the state where each coupling shaft 214 is located in the middle of the guide portions 235 Y in the circumferential direction and at the radially outer end of the long hole 16 h (see FIG. 13 ).
  • each coupling shaft 214 coupling the two plate members 220 Y is pressed against the guide surfaces 236 of the guide portions 235 Y in the inertial mass body 23 Y due to a centrifugal force acting on the weight body 22 Y and rolls or slides on the guide surfaces 236 toward one ends of the guide portions 235 Y.
  • each coupling shaft 214 moves in the radial direction of the driven plate 16 Y along the long hole 16 h of the driven plate 16 Y toward the radially inner end of the long hole 16 h .
  • each weight body 22 Y reciprocates (swings) relative to the driven plate 16 Y in the radial direction within the long hole 16 h and reciprocates (swings) relative to the inertial mass body 23 Y along the guide portions 235 Y.
  • the inertial mass body 23 Y thus swings (reciprocating rotary motion) about the center of rotation RC of the first driven plate 16 with the movement (swinging) of the weight bodies 22 Y.
  • Vibration in opposite phase from vibration transmitted from the engine EG to the drive member 11 is thus applied from the swinging inertial mass body 23 to the driven plate 16 Y via each guide portion 235 Y and each coupling shaft 214 , whereby vibration of the driven plate 16 Y can be dampened.
  • L3 represents the distance between the center of gravity G of the weight body 22 Y and the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y (second guide portions)
  • L4 represents the distance between the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y and the center of rotation RC (see FIG. 13 ).
  • the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y is the same as the center of curvature of the guide surfaces 236 (guide portions 235 Y).
  • L 4/( L 3+ L 4) ⁇ + ⁇ n+ ⁇ (10)
  • an equation of motion as given by the following expression (12) can be formed for the driven member 15 that is one of objects whose vibration is to be damped, and an equation of motion as given by the following expression (13) can be formed for the entire vibration damping device 20 Y.
  • J 1 represents the moment of inertia of the driven member 15 (first and second driven plates 16 , 17 )
  • J 2 represents the moment of inertia of the inertial mass body 23
  • m represents the mass (sum) of all (four in the present embodiment) of the weight bodies 22 Y
  • k represents stiffness (spring constant) of a system from the engine EG to the driven member 15 .
  • L 3 represents the distance between the center of gravity G of the weight body 22 Y and the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y (second guide portions)
  • L 4 represents the distance between the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y and the center of rotation RC
  • r represents the distance between the center of gravity G of the weight body 22 Y and the center of rotation RC.
  • ⁇ represents the angular velocity of steady rotation of the driven member 15
  • ⁇ represents angular displacement of the driven member 15 from the steady rotation state
  • ⁇ represents the angle formed by a line segment from the center of rotation RC to the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y and a line segment from the center of rotation RC to the center of gravity G of the weight body 22 Y
  • ⁇ represents the angle formed by a straight line passing through the center of rotation RC and the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y and a line segment from the fulcrum of swinging of the weight body 22 Y along the guide portions 235 Y to the center of gravity G of the weight body 22 Y.
  • Td represents the torque that is transferred from the engine EG to the driven member 15
  • ⁇ represents the frequency of vibration that is transmitted from the engine EG to the driven member 15
  • the symbol “′” represents the first derivative with respect to the angle ⁇
  • the symbol “′′” represents the second derivative with respect to the angle ⁇ .
  • a linearized equation of motion for the driven member 15 as given by the following expression (14) can be obtained and a linearized equation of motion for the entire vibration damping device 20 Y as given by the following expression (15) can be obtained by linearizing the expressions (12) and (13) with the angle ⁇ and the angle ⁇ being very small.
  • the angular displacement ⁇ of the driven member 15 from the steady rotation state can be regarded as zero near the equilibrium state. Accordingly, the natural angular frequency ⁇ a of the vibration damping device 20 Y near the equilibrium state as given by the above expression (8) can be obtained by ignoring the first term on the left side of the expression (15).
  • the reference order q ref namely the value to which the effective order q eff converges as the amplitude ⁇ of vibration of input torque decreases, as given by the above expression (9) can be obtained from the expression (8).
  • the vibration damping device 20 Y can be designed so that the reference order q ref is larger than the excitation order q tag of the engine EG.
  • the design values of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m which make the reference order q ref larger than the excitation order q tag and make the peak value of the torque fluctuation T Fluc closer to the minimal value MIN can be obtained by performing processing similar to that shown in FIG. 8 or 9 .
  • the reference order q ref of the vibration damping device 20 is determined according to the design values of the moments of inertia J 1 , J 2 , the axis-to-axis distances L3, L4, and the mass m thus determined.
  • the vibration damping device 20 Y may have a plurality of cylindrical outer rings 27 Y each rotatably supported by the coupling shaft 214 via a plurality of rollers (or balls, namely rolling elements) 26 Y and forming a rolling bearing.
  • three outer rings 27 Y are attached to each coupling shaft 214 so as to roll or slide on the inner surface of the long hole 16 h of the driven plate 16 Y and the guide portions 235 Y (guide surfaces 236 ) of the inertial mass body 23 Y (annular members 230 Y). This allows each weight body 22 Y and the inertial mass body 23 Y to swing more smoothly.
  • the guide surface 236 of the guide portion 235 Y is a concave cylindrical surface with a constant radius of curvature.
  • the guide surface 236 may be a concave surface whose radius of curvature varies stepwise or gradually.
  • the support surface 237 and the stopper surfaces 238 may be omitted from the guide portion 235 Y.
  • the inertial mass body 23 Y need not necessarily be supported by the driven plate 16 Y so that the inertial mass body 23 Y can rotate about the center of rotation RC.
  • the inertial mass body 23 can be made to swing symmetrically by forming the long holes 16 h in the driven plate 16 Y so that the central axis of each long hole 16 h extends in the radial direction of the driven plate 16 Y and passes through the center of rotation RC.
  • the present disclosure is not limited to this. That is, as shown in FIG. 16 , the long holes 16 h may be formed in the driven plate 16 Y so that the central axis of each long hole 16 h has the shape of a circular arc. In this case, as shown in FIG.
  • the vibration damping device 20 Y can be made to operate similarly to the vibration damping device 20 by placing the center of curvature of the central axis of each long hole 16 h on the central axis of the first coupling shaft 21 in the vibration damping device 20 and making the radius of curvature of the central axis of each long hole 16 h equal to the axis-to-axis distance L2 between the first coupling shaft 21 and the second coupling shaft 24 in the vibration damping device 20 .
  • the vibration damping device 20 Y that is a slider crank mechanism may include two driven plates 16 Y serving as support members, an inertial mass body 23 Y that is a single annular member disposed between the two driven plates 16 Y in the axial direction, and a plurality of weight bodies 22 Y each guided by long holes 16 h in each driven plate 16 Y and a guide portion 235 (guide surface 236 ) of the inertial mass body 23 Y.
  • the weight body 22 Y may include a large diameter body 22 a that is guided by the guide portion 235 Y of the inertial mass body 23 Y and shaft portions 22 b extended to both sides in the axial direction from the body 22 a so that each shaft portion 22 b is guided by the long hole 16 h in a corresponding one of the driven plates 16 Y.
  • guide portions (second guide portions) corresponding to the guide portions 235 Y may be formed in the weight bodies 22 Y, and the coupling shafts 214 may be coupled (fixed) to the inertial mass body 23 Y.
  • first guide portions corresponding to the long holes 16 h may be formed in the weight bodies 22 Y.
  • the second guide portions corresponding to the guide portions 235 Y may be formed in either the driven plates 16 Y (support members) or the weight bodies 22 Y, and the coupling shafts 214 may be disposed in either the driven plates 16 Y or the weight bodies 22 Y where the second guide portions are not formed.
  • the first guide portions corresponding to the long holes 16 h may be formed in the inertial mass body 23 Y.
  • the second guide portions corresponding to the guide portions 235 Y may be formed in either the driven plates 16 Y or the weight bodies 22 Y, and the coupling shafts 214 may be disposed in either the driven plates 16 Y or the weight bodies 22 Y where the second guide portions are not formed.
  • the vibration damping device 20 , 20 X, 20 Y may be coupled to the intermediate member 12 of the damper device 10 or may be coupled to the drive member (input element) 11 (see long dashed double-short dashed lines in FIG. 1 ).
  • the vibration damping device 20 , 20 X, 20 Y may be applied to a damper device 10 B shown in FIG. 18 .
  • the damper device 10 B of FIG. 18 corresponds to the damper device 10 having the intermediate member 12 omitted therefrom.
  • the damper device 10 B includes a drive member (input element) 11 and a driven member 15 (output element) as rotary elements and includes springs SP disposed between the drive member 11 and the driven member 15 as torque transfer elements.
  • the vibration damping device 20 , 20 X, 20 Y may be coupled to the driven member 15 of the damper device 10 B as shown in the figure or may be coupled to the drive member 11 as shown by long dashed double-short dashed lines in the figure.
  • the vibration damping device 20 , 20 X, 20 Y may be applied to a damper device 10 C shown in FIG. 19 .
  • the damper device 10 C of FIG. 19 includes a drive member (input element) 11 , a first intermediate member (first intermediate element) 121 , a second intermediate member (second intermediate element) 122 , and a driven member (output element) 15 as rotary elements, and includes first springs SP 1 disposed between the drive member 11 and the first intermediate member 121 , second springs SP 2 disposed between the second intermediate member 122 and the driven member 15 , and third springs SP 3 disposed between the first intermediate member 121 and the second intermediate member 122 as torque transfer elements.
  • the vibration damping device 20 , 20 X, 20 Y may be coupled to the driven member 15 of the damper device 10 C as shown in the figure, or may be coupled to the first intermediate member 121 , the second intermediate member 122 , or the drive member 11 as shown by long dashed double-short dashed lines in the figure.
  • coupling the vibration damping device 20 , 20 X, 20 Y to the rotary element of the damper device 10 , 10 B, 10 C allows vibration to be very satisfactorily damped by both the damper device 10 to 10 C and the vibration damping device 20 , 20 X, 20 Y while restraining an increase in weight of the damper device 10 to 10 C.
  • a vibration damping device of the present disclosure is a vibration damping device ( 20 , 20 X, 20 Y) including a support member ( 16 , 16 Y) that rotates, together with a rotary element ( 15 ) to which torque from an engine (EG) is transferred, about a center of rotation (RC) of the rotary element ( 15 ), a restoring force generating member ( 22 , 22 Y) that is coupled to the support member ( 16 , 16 Y) and that is swingable as the support member ( 16 , 16 Y) rotates, and an inertial mass body ( 23 , 23 X, 23 Y) that is coupled to the support member ( 16 , 16 Y) via the restoring force generating member ( 22 , 22 Y) and that swings, with motion of the restoring force generating member ( 22 , 22 Y), about the center of rotation (RC) as the support member ( 16 , 16 Y) rotates, wherein moments of inertia J 1 , J 2 ,
  • vibration damping capability can be improved by making a reference order, namely a value to which an order of the vibration damping device converges as amplitude of vibration of input torque decreases, larger than an excitation order of the engine rather than by making the reference order equal to the excitation order.
  • the moments of inertia J 1 , J 2 , the mass m, and the distances L 3 and L 4 are determined so that the torque fluctuation of the object for which vibration is to be damped, which is derived based on the angular displacement ⁇ and the angles ⁇ and ⁇ obtained by solving the above expressions (4) and (5), is equal to the target value.
  • the reference order can thus be made to fall within a proper range that is larger than the excitation order of the engine, whereby the vibration damping capability is further improved.
  • the target value may be defined as a value that is smaller than the torque fluctuation of the object at the time a reference order (q ref ), namely a convergence value of an order of the vibration damping device ( 20 , 20 X, 20 Y), is equal to an excitation order (q tag ) of the engine (EG).
  • the vibration damping device may be designed to satisfy 1.00 ⁇ q tag ⁇ q ref ⁇ 1.03 ⁇ q tag , more preferably 1.01 ⁇ q tag ⁇ q ref ⁇ 1.02 ⁇ q tag , where “q ref ” represents the reference order and “q tag ” represents the excitation order. This very satisfactorily improves the vibration damping capability of the vibration damping device including the restoring force generating member and the inertial mass body that swings with motion of the restoring force generating member.
  • the vibration damping device may be designed so that the order (q eff ) of the vibration damping device increases or does not vary as amplitude ( ⁇ ) of vibration of the input torque increases. Accordingly, even if any deviation of the order of the vibration damping device occurs as the amplitude of the vibration of the input torque increases, degradation in vibration damping capability due to the deviation of the order is restrained from occurring in a low engine speed range. The vibration damping capability in the low engine speed range is thus further improved.
  • the vibration damping device ( 20 ) may further include: a first coupling shaft ( 21 ) that couples the support member ( 16 ) and the restoring force generating member ( 22 ) so as to allow the support member ( 16 ) and the restoring force generating member ( 22 ) to rotate relative to each other; a second coupling shaft ( 24 ) that is supported by one of the restoring force generating member ( 22 ) and the inertial mass body ( 23 ) and that couples the restoring force generating member ( 22 ) and the inertial mass body ( 23 ) so as to allow the restoring force generating member ( 22 ) and the inertial mass body ( 23 ) to rotate relative to each other; and a guide portion ( 235 , 235 X) that is formed in the other of the restoring force generating member ( 22 ) and the inertial mass body ( 23 ) and that, as the support member ( 16 ) rotates, guides the second coupling shaft ( 24 ) so that the second coupling shaft ( 24 ) swings about the first
  • the vibration damping device ( 20 X) may further include: a connecting member ( 35 ) that is rotatably coupled to the restoring force generating member ( 22 ) via a second coupling shaft ( 24 ) and is rotatably coupled to the inertial mass body ( 23 X) via a third coupling shaft ( 30 ).
  • the vibration damping device ( 20 , 20 X) may satisfy L1+L2>L 3 +L4, where “L1” represents an axis-to-axis distance between the center of rotation (RC) of the rotary element ( 15 ) and the first coupling shaft ( 21 ), “L2” represents the axis-to-axis distance between the first coupling shaft ( 21 ) and the second coupling shaft ( 24 , 24 X), “L 3 ” represents the axis-to-axis distance between the second coupling shaft ( 24 , 24 X) and the third coupling shaft ( 25 , 30 ), and “L4” represents an axis-to-axis distance between the third coupling shaft ( 25 , 30 ) and the center of rotation (RC).
  • the vibration damping device ( 20 Y) may further include: a first guide portion ( 16 h ) that is formed in one of the support member ( 16 Y), the restoring force generating member ( 22 Y), and the inertial mass body ( 23 Y) and extends in the radial direction of the support member ( 16 Y); and a second guide portion ( 235 Y) that is formed in one of the two other than the one of the support member ( 16 Y), the restoring force generating member ( 22 Y), and the inertial mass body ( 23 Y) and has a shape of a circular arc, wherein the other of the two other than the one of the support member ( 16 Y), the restoring force generating member ( 22 Y), and the inertial mass body ( 23 Y) may be guided by the first and second guide portions ( 16 h , 235 Y).
  • this vibration damping device by increasing the order of the vibration damping device as the amplitude of vibration of input torque increases, the vibration damping capability is further improved while
  • the support member ( 16 , 16 Y) may rotate coaxially and together with one of rotary elements of a damper device ( 10 , 10 B, 10 C) having a plurality of rotary elements ( 11 , 12 , 121 , 122 , 15 ) including at least an input element ( 11 ) and an output element ( 15 ) and an elastic body (SP, SP 1 , SP 2 , SP 3 ) that transfers the torque between the input element ( 11 ) and the output element ( 15 ). Coupling the vibration damping device to the rotary element of the damper device in this manner allows vibration to be very satisfactorily damped by both the damper device and the vibration damping device while restraining an increase in weight of the damper device.
  • the output element ( 15 ) of the damper device ( 10 , 10 B, 10 C) may be operatively (directly or indirectly) coupled to an input shaft (IS) of a transmission (TM).
  • a method for designing a vibration damping device is a method for designing a vibration damping device ( 20 , 20 X, 20 Y) including a support member ( 16 , 16 Y) that rotates, together with a rotary element ( 15 ) to which torque from an engine (EG) is transferred, about a center of rotation (RC) of the rotary element ( 15 ), a restoring force generating member ( 22 , 22 Y) that is coupled to the support member ( 16 , 16 Y) and that is swingable as the support member ( 16 , 16 Y) rotates, and an inertial mass body ( 23 , 23 X, 23 Y) that is coupled to the support member ( 16 , 16 Y) via the restoring force generating member ( 22 , 22 Y) and that swings, with motion of the restoring force generating member ( 22 , 22 Y), about the center of rotation (RC) as the support member ( 16 , 16 Y) rotates.
  • the method comprises determining moments of inertia J 1 , J 2 , mass m, and distances L 3 and L 4 so that torque fluctuation of an object for which vibration is to be damped, which is derived based on angular displacement ⁇ and angles ⁇ and ⁇ obtained by solving the above expressions (4) and (5), is equal to a target value.
  • Designing the vibration damping device by this method allows a reference order, namely a convergence value of an order of the vibration damping device, to fall within a proper range that is larger than an excitation order of the engine and thus further improves vibration damping capability.
  • the target value may be defined as a value that is smaller than the torque fluctuation of the object at the time a reference order (q ref ), namely a convergence value of an order of the vibration damping device ( 20 , 20 X, 20 Y), is equal to an excitation order (q tag ) of the engine (EG).
  • the torque fluctuation may be derived by changing the moments of inertias J 1 , J 2 , the mass m, and the distances L 3 and L 4 so that the reference order (q ref ), namely the convergence value of the order of the vibration damping device ( 20 , 20 X, 20 Y), increases, and the moments of inertias J 1 , J 2 , the mass m, and the distances L 3 and L4 which make a peak value of the torque fluctuation in a predetermined rotational speed range equal to a minimal value or to a value close to the minimal value may be used as design values.
  • the reference order (q ref ) namely the convergence value of the order of the vibration damping device ( 20 , 20 X, 20 Y)
  • the moments of inertias J 1 , J 2 , the mass m, and the distances L 3 and L4 which make a peak value of the torque fluctuation in a predetermined rotational speed range equal to a minimal value or to a value

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  • Mechanical Engineering (AREA)
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  • Aviation & Aerospace Engineering (AREA)
  • Vibration Prevention Devices (AREA)
  • Mechanical Operated Clutches (AREA)
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